Arrangement including a gear pump

ABSTRACT

An arrangement including a gear pump ( 1 ), comprising a pump housing ( 10 ), two meshing gear wheels ( 11, 12 ) contained in the pump housing ( 10 ) and two shafts ( 2, 3 ) which are operatively connected to the gear wheels ( 11, 12 ) and extend through the pump housing ( 10 ), is disclosed wherein the two shafts ( 2, 3 ) are each operatively connected to respective drive units ( 7, 8 ). The arrangement includes a coupling unit ( 22 ) for compensation of eccentricities between the drive unit ( 7, 8 ) and the respective shaft ( 2, 3 ) is arranged between each gear wheel ( 11, 12 ) and drive unit ( 7, 8 ) and that a rotary encoder/sensor unit ( 24, 25 ) is arranged between the center of the gear wheel and the center of the respective drive.

TECHNICAL FIELD

The present invention relates to an arrangement including a gear pump,comprising a pump housing, two meshing gear wheels contained in the pumphousing and two shafts which are operatively connected to the gearwheels and extend through the pump housing, wherein the two shafts areeach operatively connected to respective drive units.

BACKGROUND AND SUMMARY

Gear pumps consist of two meshing gear wheels which are mounted onshafts, wherein generally one shaft is connected to a drive unit. Theshaft which is not being driven by a drive unit is driven by torquetransmission from the shaft being driven via the tooth flanks.

Often problems of wear occur at the tooth flanks through excessivecontact pressure due to the torque transmission. Namely, on the one handbecause the pressure load brought about by the pressure difference needsto be transferred via the tooth flanks from the shaft being driven tothe driven shaft, and on the other hand because the friction needs to beovercome. Damages on the tooth flanks can occur through abrasion or wear(pittings, micro-weldings, abrasion), especially in the manufacturing ofpolymers in large polymerization installations or in compounding ofplastics with large throughputs at very high backpressures and hightemperatures, i.e. overall high torque.

In order to avoid these damages two shaft drives have already been used,for which the propulsion takes place with a single drive unit (motor),and then subsequently the force is distributed by a mechanical transferbox to the two gear pump shafts.

Furthermore, from the patent CH-659 290 a gear pump is known in whichthe two shafts are each driven with a drive unit. Each of the two gearwheels draws the necessary drive power from the associated drive unit.Only relatively minor power differences are transmitted between the twogear wheels.

From the EP-0 886 068 B1 a gear pump is known in which again two driveunits are provided for individually driving the shafts, where the phaseand angular velocity of the meshing gear wheels are coordinated in sucha way that on the one hand a lifting off of tooth flanks of the meshinggear wheels and on the other hand an excessive excess torque on thetooth flanks of the meshing gear wheels are avoided.

It has been shown that in known gear pumps the wear, especially whenconveying abrasive media, can be substantial.

In particular in extrusion applications of highly filled, abrasivepolymer melts the problem of high tooth flank wear due to abrasion canoccur and thus the problem of premature failure of the gear wheel shaftsarises, since abrasive particles contained in the melt are grindedbetween the tooth flanks. Thereby, damage and abrasion of the surface ofthe tooth flanks can be caused. Furthermore, the viscosity of the meltis increased due to loading with filler material and therewith thetorque requirement of the entire pump or the required torque at theindividual shafts rises, so that a possible exceedance of the allowedcontact pressure at the tooth flanks again becomes the focus ofattention.

It is therefore an object of the present invention to provide anarrangement including a gear pump with which an improvement is achievedwith respect to a least one of the mentioned disadvantages.

This object is achieved by the improved arrangement of the inventionwherein a coupling unit for compensation of eccentricities between thedrive unit and the respective shaft is arranged between each gear wheeland drive unit and wherein a rotary encoder/sensor unit is arrangedbetween the center of the gear wheel and the center of the respectivedrive. Preferred embodiments are disclosed hereinafter.

The present invention relates to an arrangement including a gear pump,comprising a pump housing, two meshing gear wheels contained in the pumphousing and two shafts which are operatively connected to the gearwheels and extend through the pump housing, wherein the two shafts areeach operatively connected to respective drive units. According to theinvention it is provided that a coupling unit for compensation ofeccentricities between the drive unit and the respective shaft isarranged between each gear wheel and drive unit and that a rotaryencoder/sensor unit is arranged between the center of the gear wheel andthe center of the respective drive.

One embodiment of the arrangement according to the invention ischaracterized in that the rotary encoder/sensor unit is located in anaxial region which is defined by the center between the center of thegear wheel and the center of the drive plus a deviation on both sides ofat most 10% of the distance between the center of the gear wheel and thecenter of the drive.

Further embodiments of the arrangement according to the invention arecharacterized in that rotary encoders/sensor units are respectivelyarranged in the middle between the respective center of the gear wheeland the respective center of the drive.

Further embodiments of the arrangement according to the invention arecharacterized in that the rotary encoders/sensor units feature a radialdistance to the rotation axis of the respective shaft which is larger,preferably at least twice as large, as the outer radius of the gearwheels.

Further embodiments of the arrangement according to the invention arecharacterized in that the rotary encoders/sensor units are eitheroptical or magnetic rotary encoders/sensor units.

Further embodiments of the arrangement according to the invention arecharacterized in that the rotary encoders/sensor units are arranged suchthat a connecting line which runs through the corresponding rotaryencoder/sensor unit and extends perpendicularly from the shaft enclosestogether with a plane which extends centrally between the two rotationaxes on a suction side an angle in the range of 35° to 55°, preferably40° to 50°, preferably 45°.

Further embodiments of the arrangement according to the invention arecharacterized in that each drive unit features a rotor and a stator,wherein the rotor is axially moveable with respect to the stator.

Further embodiments of the arrangement according to the invention arecharacterized in that the drive unit features on the far side withrespect to the gear pump a differential bearing unit which radiallysupports the rotor of the drive unit.

Further embodiments of the arrangement according to the invention arecharacterized in that the rotor of the drive unit is connected to therespective shaft of the gear pump via a coupling unit.

Further embodiments of the arrangement according to the invention arecharacterized in that the coupling unit is a membrane coupling.

Further embodiments of the arrangement according to the invention arecharacterized in that a flange is arranged between the pump housing andthe stator of the respective drive unit, wherein the flange featuresbores through which a cooling agent circulates for adjusting thetemperature.

Further embodiments of the arrangement according to the invention arecharacterized in that the drive units are connectable to the respectiveshaft of the gear pump from the far side with respect to the gear pump.

Further embodiments of the arrangement according to the invention arecharacterized in that the connections between the drive units and therespective shafts of the gear pump are conical polygon connections.

Further embodiments of the arrangement according to the invention arecharacterized in that the drive units are of the torque motor type.

Further embodiments of the arrangement according to the invention arecharacterized in that the one drive unit, the gear pump and the otherdrive unit are each contained in a temperature zone in which thetemperatures are adjustable to specified values, wherein preferablyisolation regions are present between neighboring temperature zones.

Further embodiments of the arrangement according to the invention arecharacterized in that a current position of the one gear wheel isdeterminable with respect to the current position of the other gearwheel and that the current position of the one gear wheel iscontinuously adjustable with respect to the current position of theother gear wheel according to specified predefined operating conditions.

Further embodiments of the arrangement according to the invention arecharacterized in that the determination of the present position of theone gear wheel with respect to the present position of the other gearwheel is adjustable via a reference value, which is determinable beforethe normal operation of the gear pump or during interruptions of thenormal operation of the gear pump.

Further embodiments of the arrangement according to the invention arecharacterized in that the reference value lies in the middle betweentooth flanks of a tooth space of a gear wheel, preferably in the middlebetween tooth flanks of a tooth space of a gear wheel.

Further embodiments of the arrangement according to the invention arecharacterized in that the reference value can be determined in that,

-   -   the one gear wheel can be driven by the other gear wheel with a        predetermined torque,    -   a first angle difference can be determined by calculating the        difference between values measured with the rotary        encoders/sensor units,    -   the other gear wheel can be driven by the one gear wheel with a        predetermined torque,    -   a second angle difference can be determined by calculating the        difference between values measured with the rotary        encoders/sensor units,    -   a difference can be determined between the first angle        difference and the second angle difference, and    -   the reference value can be specified within the determined        difference.

Thus an arrangement including a gear pump is provided which can beautomatically calibrated. The arrangement can perform this calibrationboth before the beginning of operation as well as during serviceinterruptions without further action by the operator. With thisarrangement a possible wear of tooth flanks can also be detected, sincethen the difference between the first angle difference and the secondangle difference also changes or increases. An excessive wear can thenbe simply detected when exceeding a threshold value.

Further embodiments of the arrangement according to the invention arecharacterized in that at least one of the following current values canbe monitored:

-   -   the first angle difference,    -   a difference between the first angle difference and the        reference value,    -   the second angle difference,    -   a difference between the second angle difference and the        reference value,        and that in case of under- or overshooting of the at least one        of the current values below a predefined value at least one of        the following actions can be executed:    -   an optical warning,    -   optical display,    -   acoustic warning,    -   change of operating conditions of the gear pump.

Further embodiments of the arrangement according to the invention arecharacterized in that rotary encoders/sensor units can be applied fordetermining the current position of the one gear wheel and the secondgear wheel, wherein each rotary encoder/sensor unit is arrangedcentrally between the toothing of the respective gear wheel and a rotorof the respective drive unit.

A central arrangement of the rotary encoders/sensor units has theadvantage that an existing torsion angle due to a non-ideal stiffness ofthe entire drive train has a reduced influence on the measurement errorof the system. The measurement error is halved by the centralarrangement.

Further embodiments of the arrangement according to the invention arecharacterized in that a predefined gear lash can be adjusted between twomeshing gears.

Further embodiments of the arrangement according to the invention arecharacterized in that a leading flank, in the direction of rotation ofthe one gear wheel, of a tooth meshing into a tooth gap touches alagging flank, in the direction of rotation of the other gear wheel, andthat a leading flank, in the direction of rotation of the other gearwheel, of a tooth meshing into a tooth gap touches a lagging flank, inthe direction of rotation of the one gear wheel.

This operation condition is also referred to a changeover of flankssince the flanks touching each another will change during the course ofthe extrusion process.

Further embodiments of the arrangement according to the invention arecharacterized in that the one gear wheel drives the other gear wheelwith a predetermined torque, wherein the predetermined torque is greaterthan half of the total torque generated by both drives.

A precise torque setting is achievable by appropriate control of therotary speed or current positions of the gear wheels to each other. Thetooth flanks thus transfer an arbitrary adjustable torque, however theynever lift off from each other during operation if a defined flanksealing is always to be achieved.

Further embodiments of the arrangement according to the invention arecharacterized in that the rotary speed of the shafts driven by the driveunits can be synchronously adjusted in such a way that a pressure of themedium to be conveyed stays substantially constant on a discharge sideof the gear pump.

The advantage associate with this is that no disturbing pulsation ispresent at the discharge side of the gear pump anymore, which isreflected in the quality of the extrudate.

Further embodiments of the arrangement according to the invention arecharacterized in that the pressure of the medium to be conveyed ismeasured on the discharge side of the gear pump and that the rotaryspeed can be adjusted in dependence of the measured pressure of themedium.

BRIEF DESCRIPTION OF DRAWINGS

The invention will be explained in the following with the help ofdrawings in which embodiments of the present invention are illustrated.Thereby showing:

FIG. 1 a known arrangement with one gear pump and one drive unit,

FIG. 2 a sectional view through the cutting plane A-A indicated in FIG.4 of an arrangement according to the invention including a gear pump anda drive unit,

FIG. 3 a schematic representation of the arrangement according to theinvention with information on temperature zones,

FIG. 4 a position for rotary encoder and sensor unit to determine thecurrent position of the gear wheels,

FIG. 5 a transversal cut through the rotation axes of the shafts in areaof the gear wheels to illustrate a first operating condition,

FIG. 6 again a transversal cut through the rotation axes of the shaftsin area of the gear wheels to illustrate a second operating condition,

FIG. 7 again a transversal cut through the rotation axes of the shaftsin area of the gear wheels to illustrate a third operating condition,

FIG. 8 again a transversal cut through the rotation axes of the shaftsin area of the gear wheels to illustrate a fourth operating condition,and

FIG. 9 a graph with a rotary speed curve, a pressure curve and a torquecurve as a function of time.

DETAILED DESCRIPTION

In FIG. 1 a known arrangement including a gear pump 1 is shown whichconveys a medium F to be conveyed from a suction side S to a dischargeside D. A pump housing 10 can be seen in FIG. 1 through which two shafts2 and 3 extend to the outside. The shaft 3 extending to the outside isconnected to a drive unit 7 via a first universal joint 4, an axlesegment 6 whose length is adjustable and a second universal joint 5.Accordingly, also the shaft 2 extending to the outside is connected to afurther drive unit (not shown in FIG. 1) via a corresponding first andsecond universal joint as well as via a corresponding axle segment.Thus, the gear wheels (not visible in FIG. 1) are each propelled by anown drive unit.

It should be noted that the double universal joint consisting of thefirst and the second universal joint 4 and 5 together with theadjustable axle segment 6 is provided to accommodate lateral and angulardeviations of the drive unit in relation to the shafts 2 or 3. Throughthe double universal joint in combination with the adjustable axlesegment 6 an additional bearing force acts on a shaft bearing containedin the pump housing 10. This additional bearing force arises due to theself-weight of the double universal joint and the axle segment 6. Theadditional bearing force is considerable because of a relatively shortdistance between the pump bearings, which are located in the pumphousing 10 to support the shafts 2 and 3, with respect to the length ofthe double universal joint.

FIG. 2 shows a sectional view through an arrangement according to theinvention including a gear pump 1, wherein the cutting plane is placedin the rotary axes 13 and 14 of the shafts 2 and 3 and through a sensor25, according to the cutting plane A-A marked in FIG. 4. For the sake ofsimplicity FIG. 2 only shows half of the gear pump 1. Accordingly, alsoonly one drive unit 7 is shown. The drive unit 7 is pressed via a flange15 directly, i.e. without an intermediate gearing, to the pump housing10 or rather its cover. The rotating parts of the drive unit 7, such asa hub 16, membrane coupling 22 and a rotor 18, are connected to theshaft 3 of the gear pump 1 via a screw 21. The screw 21 can be loosenedif required, whereby the drive unit 7 in turn can be unfastened from thegear pump 1. After the loosening of the screw 40, which joins the flange15 with the pump housing 10 or rather its cover, and after loosening thescrew 21 the entire drive unit 7 can be unfastened from the gear pump 1.The shafts 2, 3 of the gear pump 1 and their bearing unit remain withinthe gear pump 1 and can be disassembled individually.

Apart from the flange 15 and the hub 16, the drive unit 7 furthercomprises a rotor 18, a stator 17 and a drive cover 19 with an opening20. The drive cover 19 closes the drive unit 7 on the far side of thegear pump 1 and is connected to the stator 17, wherein the opening 20 iscentrally arranged on the extended rotary axis 13 of the shaft 3. On thenear side of the gear pump 1, the stator 17 is connected to the flange15.

As already stated, the gear pump 1 is connected directly, i.e. withoutan intermediate gearing, to the drive unit 7. For this the screw 21 isprovided, with the help of which the rotor 18 is axially fixed via thehub 16 and the flange 15. During the assembly of the drive unit 7 to thegear pump 1 the screw 21 is passed through the opening 20 in the driveunit 7 and along the rotary axis 13 of the shaft 3 and is fastened in acorresponding bore in the shaft 3. Thereby, the hub 16 is connected tothe shaft 3 via a so-called conical polygon connection, which on the onehand makes possible an exact alignment of the rotor 18 with the shaft 3,and on the other hand makes possible an extremely torsion proofconnection between the rotor 18, the drive unit 7 and the shaft 3, whichis to be driven, of the gear pump 1.

It is already clear from the comparison of the known arrangementaccording to FIG. 1 and the arrangement according to the inventionaccording to FIG. 2 that the arrangement according to the invention iscomparatively extremely short and also yields a torsion proof connectiondue to the short rotary axis connection between the rotor 18 and thegear wheel 11. This is of importance especially in connection with themethod according to the invention that is still to be explained.

Since, as with the double universal joint according to FIG. 1, thearrangement according to FIG. 2 also requires an angular and a lateralcompensation, on the one hand a membrane coupling 22 at the side of thegear pump end of the rotor 18—for the angular compensation—and on theother hand the stator 17 and the rotor 18 are formed such that the rotor18 can be axially moved with respect to the stator 17, in order to makepossible a lateral compensation.

It is for instance also conceivable that the membrane coupling 22 andthe hub 16 are a single part, as apparent from FIG. 2, wherein in theleft half on the driving side the single part fulfils the classicalfunction of a hub, which can be coupled to the shaft 3, and wherein inthe right half this single part is thin-walled and fulfils the functionof the membrane coupling.

Apart from the stated support of the rotor 18 with respect to the stator17 on the side of the gear pump 1 by means of flange 15 and hub 16—viaconical polygon connection and screw 21—, a differential bearing unit 23is provided on the far side with respect to the gear pump 1, thatradially holds the rotor in position with respect to the stator 17.

Eccentricities of the rotary axis 13 of the shaft 3 with respect to arotary axis of the differential bearing 23 due to manufacturingtolerances can be compensated for via a membrane coupling 22. Indeed anadditional loading of the gear pump bearings arises because of theeccentricities due to manufacturing tolerances, however the resultingreactions of the moment remain within relatively narrow boundaries,since the distance of the membrane coupling 22 to the loaded bearing isshort and since only a moderate angular compensation has to beaccomplished.

For instance, a so-called torque motor is employed as drive unit 7,which is a multi-pole, permanently excited three-phase synchronous motorwith hollow rotor shaft, for the direct coupling to the gear pump 1stated above. Torque motors are characterized in particular by a shortcompact design and a low skew slackness (high torsional stiffness).

As will become apparent from the following explanations regarding theoperation of the arrangement including a gear pump 1, accurateinformation on the current position of the one gear wheel with respectto the other gear wheel is important. At the same time, there is ademand for a direct and unbiased influence of the drive units on thegear wheels of the gear pump. One criterion is the already stated lowskew slackness (high stiffness) between the drive unit and the drivengear wheel. A further criterion is an as accurate as possiblemeasurement of the current position of the one gear wheel with respectto the current position of the other gear wheel.

In the embodiment according to FIG. 2 this is achieved in that a rotaryencoder 24 is arranged at the periphery of the hub 16 which interactswith a sensor unit 25 that is connected to the stator 17. For instance,a grid pattern is applied on the hub 16 that can be read by the sensorunit 25. Instead of such an optical measurement device a correspondingmagnetic measurement device or another method for determining theposition can be employed.

In order to minimize a possible measurement error due to an eccentricitybetween the rotary encoder 24 and the gear teeth, the diameter of therotary encoder 24 is implemented as large as possible. The eccentricityof the rotary encoder 24 itself is minimized by integration of thereceptacle for the rotary encoder 24 into the hub 16. Since the hub 16is made in one-piece very tight manufacturing tolerances can bemaintained for the receptacle for the rotary encoder 24.

The position of the rotary encoder 24 and of the sensor unit 25 ispreferably chosen between the middle of the rotor 18 and the stator 17and the middle of the driven gear wheel 11 of the gear pump 1. For auniform stiffness distribution over the drive train (i.e. between thecenter of the rotor 18 and stator 17 and the center of the driven gearwheel 11 of the gear pump 1) the rotary encoder 24 is preferablyarranged in the middle between the center of the rotor 18 and stator 17and the center of the driven gear wheel 11 of the gear pump 1.

One possible area of application of the arrangement including a gearpump is the pressure build-up downstream from the extruder in conveyingof plastic melts in an extrusion line. In these applications polymermelts are conveyed at a temperature of 300° C. against high dischargepressures (e.g. 300 bar). For this high drive powers and therewith alsohigh torques are necessary. Accordingly, the gear pump and the pumphousing is heated to a temperature of for instance 300° C. due to themedium to be conveyed, whilst the temperature of the drive units 7 and 8should not exceed 60° C. particularly because of the electronic circuitsused with these. To illustrate these circumstances FIG. 3 shows thearrangement including a gear pump 1 according to the invention, wherenow the gear pump 1 and the two laterally arranged drive units 7 and 8are represented as simple blocks. The individual components arecontained in temperature zones 32, 33 and 34, which have to exhibitpermitted or required temperature values according to the foregoingstatements. Thus, the gear pump 1 is included in the temperature zone33, which is operated at the temperature of the medium to be conveyed,for instance at 300° C. On the drive side, the drive units 7 and 8 areprovided in the temperature zones 32 and 34, respectively, the maximumvalue of which is not allowed to exceed 60° C. for a proper functioning.The present arrangement requires the positioning of electricalcomponents in the immediate vicinity of the gear pump 1. Since the gearpump is heated up to 300° C. insulating separating walls 30 and 31 arerequired which are present between the temperature zones 32 and 33 andbetween the temperature zones 33 and 34, respectively. Apart from theinsulating separating walls 30 and 31 further measures are required asnecessary so that the temperature in the cold temperature zones 32 and34 does not reach inadmissible values. An additional measure forinstance consists in providing an active cooling system (e.g. an activewater cooling system).

It is also conceivable to protect the rotor 18 (FIG. 2) from overheatingby inserting a cooling of the flange 15 between the hub 16 and the gearpump 1. Thereby, the cooling is for instance implemented as star shapedbores in the flange 15. Herewith, very good cooling properties areachieved since the deflections generate high turbulences. The hub iscooled on the entire flange side surface by irradiation and forcedconvection.

FIG. 4 shows a possible positioning of the sensor unit 25, which isemployed to determine the current position of the one gear wheel withrespect to the other gear wheel, where FIG. 4 shows a cut across therotary axes 11 and 13 of the shafts 2 and 3. The medium to be conveyedis advanced in the direction of the arrow from the suction side S to thedischarge side D. In doing so a force component is generated in thedirection of the arrows P, P′, which act on the shaft bearing of thegear pump and lead to a minor displacement of the shafts 2 and 3 (FIG.2).

In order to compensate for the eccentricity caused by the displacementthe sensor unit 25 will now be arranged in the direction of thedisplacement, i.e. in the direction of the deflection of the shaft. Themounting takes place for instance at 45° and is therefore in the averageof the possible displacement angles, which is dependent on the pressuredifference and the viscosity. In case of insufficient accuracy caused bythe displacement of the shaft and the deflection, an arrangement of thesensor unit 25 can for instance be employed which is dependent on thewidth of the gear wheels and the value of the backlash.

Based on the FIGS. 5 to 9 different operating conditions are explainedin the following which can be specified as predefined procedures for theoperation of the arrangement including a gear pump.

FIG. 5 shows a cut across the rotary axes 11 and 13 in the region of thegear wheels 11 and 12. The medium F to be conveyed is picked up by thetooth gaps on the suction side S and subsequently transported along thepump housing to the discharge side D where the medium F is extruded bythe cogging gear wheels 11, 12.

During operation of the gear pump a “trapped volume” is formed in theregion of the toothing between the bottom and the face of the tooth thatis sealed off by the tooth flanks, which are almost touching each other,in front of and behind this volume. However, for fluidic purposes a flowgap can specifically be generated at those locations at which a largeflow gap is desired for tribological reasons (optimal gap width inrelation to the relative speed of the tooth flanks). The ratio of thesetwo sealing gaps can be actively controlled by the existing positioncontrol of the shafts. Once the gap moving in advance of the “trappedvolume” can be minimized and once the gap following the “trapped volume”can be minimized. Thus, it is possible to actively influence theextruding process from this “trapped volume”, with which the uniformityof the flow can be optimized.

In order to adjust the various operating states and conditions of thearrangement according to the invention information regarding the currentposition of the one gear wheel 11 with respect to the current positionof the other gear wheel 12 must be known. This information constitutesthe actual initial conditions which are necessary for further adjustmentof the gear wheels to each other. One possibility to determine thisinformation consists of the execution of the following process stepswhich constitutes a calibration:

In a first step the first shaft 2 drives the second shaft 3 with apredefined torque. Thereby, a first absolute rotation angle differenceis determined with the help of the stated rotary encoder 24 incombination with the sensor unit 25 (FIG. 2) at each shaft 2 and 3 inthat a difference is determined between the measured value of the onesensor unit 25 and the measured value of the other sensor unit 25′.

In a second step the second shaft 3 drives the first shaft 2 with thesame defined torque as in the first step. Thereby, a second absoluterotation angle difference is determined again with the help of thestated rotary encoder 24 in combination with the sensor unit 25 at eachshaft 2 and 3 in that again a difference is determined between themeasured value of the one sensor unit 25 and the measured value of theother sensor unit 25′.

In a third step the difference between the first absolute rotationdifference and the second absolute rotation difference is determined.This difference is the actual range within which the gear wheels canmove to each other, provided that the defined torque that was used inthe first and second step for the measurement is not exceeded. In thisrange a reference value can now be specified with respect to which thecurrent positions of the gear wheels are indicated. The reference valueis then an origin of a defined coordinate system. For instance thereference value lies in the middle between the tooth flanks of a toothgap such that the absolute values of the maximum displacements areidentical.

For the operating condition explained with the help of FIG. 5 a constantforce FO is transferred between the gear wheels 11 and 12 as illustratedin the detailed view X on the right hand side of FIG. 5.

In this way the operating conditions can now be chosen after thespecification of the reference point in a first setting such that a gearwheel transfers half plus a defined percentage of the total torque.Accordingly, the other gear wheel then transfers half minus the definedpercentage of the total torque.

Under these operating conditions a defined sealing can be achievedbetween the tooth flanks. The application area of these operatingconditions is aimed at conveying of low-viscosity fluids for which asealing action is necessary between the tooth flanks in order to achievea sufficient sealing between the discharge side D and the suction sideS.

A further setting consists in that the gear lash between the flanks oftwo meshing gear wheels is chosen as operating condition. Namely, forinstance in 10% steps from the contact of the flanks (no gear lash) viaa central alignment (i.e. the tooth meshing into a tooth gap liesexactly in the middle of the gap) up until the tooth flanks touch eachother again, where this time this pertains to the trailing tooth flank.

FIG. 6 illustrates the operating conditions explained above again in acut across the rotary axes 13 and 14 in the region of the gear wheels 11and 12. Here also a section X in the region of the cogging gear wheels11 and 12 is presented enlarged in detail, where further a preset gearlash 26 is highlighted.

These operating conditions are selected if the medium F to be conveyedhas a medium viscosity. With the setting of the gear lash 26 theextrusion pressure can be set so that it is preferably equal to thepressure on the discharge side D. An excessive gear lash 26, which leadsto a smaller extrusion pressure than the pressure on the discharge sideD, must be avoided, since an insufficient sealing effect is obtainedbetween the discharge side D and suction side S. The operating conditionat which a certain amount of gear lash 26 (i.e. without flank contacts)is present can then be implemented with a corrosion-resistant (and thusoften soft) coating of the gear wheels without causing damage due toabrasion.

A further setting consists in that a mode with a changeover of flanks isproposed as operating conditions. Thereby, a gear wheel changes theflanks during the theoretical roll-off of a tooth on the line of action.The extrusion pressure discharge thus always occurs toward the suctionside S.

The mode with a changeover of flanks is explained with the help of FIG.7 which again shows sectional views across the rotary axes 13, 14 in aregion of the gear wheels 11, 12. On the left hand side of FIG. 7 astate is shown in which the tooth Z₁′ of the gear wheel 11 is meshinginto a tooth gap of the gear wheel 12 and touching the tooth Z₁. On theright hand side of FIG. 7 a later state is shown in which the tooth Z₂of the gear wheel 12 is meshing into a tooth gap of the gear wheel 11and touching the tooth Z₁′. Thereby, the mentioned changeover of flankstook place in the meantime.

The mode with a changeover of flanks offers at least one of thefollowing advantages:

-   -   minimization of the pulsation due to the extrusion pressure by        means of discharging the extrusion pressure to the suction side        S;    -   minimization of the applied torque by minimization of the        extrusion pressure energy;    -   reduction of the temperature increase by minimization of the        extrusion pressure energy.

The operating conditions according to the mentioned mode with achangeover of flanks is for instance applied for highly viscous mediawith which the extrusion pressure becomes so high that a very largetorque is required in order to generate the extrusion pressure energy,since the latter represents a pure loss in terms of energy.

The extrusion behavior can be specifically varied using the flexibilityof an electronic control system in dependence of the properties of themedium to be pumped, i.e. the flow characteristics or the solid loadingof the polymer to be conveyed. In this way an optimal speed profile canbe assigned to each type of polymer.

With the help of FIG. 8 a further aspect of the method according to thepresent invention is explained. Based on knowledge about the currentposition of the one gear wheel 11 with respect to current position ofthe other gear wheel 12, for instance from application of the explainedsteps for a calibration, and the maximum backlash (difference betweenthe first absolute rotation difference and the second absolute rotationdifference) for a current position of the one gear wheel 11 with respectto the current position of the other gear wheel 12 a statement can nowbe made about the wear of the gear wheels 11, 12. For instance, when themaximum backlash for a certain torque transferred from one gear wheel11, 12 to the other gear wheel 12, 11 changes. When for example thebacklash is increased beyond a predetermined maximum threshold this canbe interpreted such that a gear wheel and/or a shaft with a gear wheelmust be replaced shortly since for instance a failure of the system mustbe expected soon. Thus, when for a gear pump set according to the lefthalf of FIG. 8 to always extrude to the suction side S the maximumallowed wear (i.e. the maximum allowed backlash on one side, departingfrom the reference value) is exceeded the operating conditions areadapted automatically or following a corresponding manual interventionof a supervising person. The adaptation of the operating conditions canthereby take place such that henceforth the necessary sealing istransferred to the other tooth flanks. According to the right half ofFIG. 8 this is the leading tooth flank of a tooth meshing into a toothgap.

It is also conceivable that when discovering a wear the current positionof the one gear wheel with respect to the current position of the othergear wheel is changed such that the originally desired optimal operatingconditions are maintained. For instance, a lifting off of the toothflanks can occur due to the wear. The corresponding correction in orderto restore the desired operating conditions would then be a change ofcurrent position of the one gear wheel with respect to the currentposition of the other gear wheel so that the tooth flanks touch eachother in the desired manner again and that the desired backlash isobtained again.

The monitoring of wear can also be utilized so that when a predeterminedwear is discovered an acoustic and/or optical warning is given to thesupervising person so that precautions can be taken to prevent a failureof the pump system. Doing so, it is conceivable that upon activation ofan alarm a replacement shaft or replacement gear wheel is ordered fromthe manufacturer early enough so that the required spare parts areavailable on-site before a possible failure of the pump system occurs.

In some extrusion systems in which gear pumps are used pressurefluctuations due to the mentioned extrusion process of residual volumesbetween the gear wheels are disturbing. These pressure fluctuations arealso termed pulsations and lead to irregularities in the productgenerated by the extrusion. For this reason, various measures have beenproposed to reduce the pressure fluctuations. To be mentioned is the useof helical gearing or double helical gearing, both of which however havesystem-related disadvantages.

According to the present invention pressure fluctuations are eliminatedor at least strongly reduced by actively influencing the rotary speed ofthe two gear wheel shafts.

The arrangement according to the invention as well as the methodaccording to the invention is able to vary the course of the rotaryspeed per extrusion process in such a way that the pressure on thedischarge side lies within narrow limits or that the pressure on thedischarge side is constant. Thus, the extrusion process of the medium tobe conveyed is specifically controlled from the bottom of the tooth viathe current position of the one gear wheel with respect to the currentposition of the other gear wheel.

It has become apparent that it is indeed generally desirable that thepressure fluctuations can be completely eliminated. However, in certainapplications specific pressure fluctuations can be desirable so thatappropriate variations are obtained in the thickness of the extrudate.Therefore the method according to the inventive opens up newmanufacturing possibilities in the field of extrusion, particularly inconnection with the arrangement according to the invention including agear pump.

In order to achieve a simple and at the same time complete eliminationof pressure fluctuations a contact ratio of 1 must be selected. If acontact ratio of 1 is chosen then always only one pair of teeth isengaged in the displacement, i.e. the extrusion (see Vogel FachbuchJarosla and Monika Ivantysyn: “Hydrostatische Pumpen and Motoren”, 1993,p. 319). In this case a sinusoidal curve results for the displacementvolume flow. It can be easily and efficiently corrected via a sinusoidalcompensation table (for instance a so-called “look-up” table).

In FIG. 9 a rotary speed curve 90 of the gear wheel pump shaft, apressure curve 91 of the pressure on the discharge side of the gear pumpand a torque curve 92 of the torque of the gear wheel pump shaft areillustrated. The rotary speed curve 90, the pressure curve 91 and thetorque curve 92 are plotted as a function of time. The rotary speed ofthe gear wheel pump shaft is adjusted in such as way as a function oftime that the pressure on the discharge side of the gear pump isconstant or at least lies within a predetermined tolerance range. Therotary speed curve 90 shown in FIG. 9 has a periodicity with a period T.It is the time period during which a meshing into a corresponding gaptakes place. If the rotary speed of both shafts is now synchronouslycontrolled according to the rotary speed curve 90 the pulsation can befully compensated.

It is pointed out that the pulsation compensation can be combined withall the operating conditions or specifications outlined in thisdescription.

Due to the periodicity it is possible to store the rotary speed curve 90in a storage unit (look-up table). The values for the rotary speed to beset are then read out in a given cycle, wherein the predetermined cyclearises on the discharge side due to the pressure to be set.

Alternatively, it is also possible, to measure the pressure on thedischarge side with a pressure sensor and to use the rotary speed basedon the measured pressure for setting the rotary speed. This so-calledon-line pressure setting procedure is indeed more costly to implementbut other applications are made possible for realizing specificmanufacturing processes in the field of extrusion.

The specific manipulation of the position control as has been explainedabove to prevent or reduce the pulsation of the pressure on thedischarge side of a gear pump can also be used to reduce the total shearloading for shear sensitive materials. For this care is taken whendetermining the rotary speed curve to ensure that a maximum shear stressis not exceeded.

The present invention makes it possible for the first time tospecifically influence the effects of pulsation, extrusion pressure andtribological behavior. With the settings all the effects, which are ofimportance for the specific case, can be taken into account, orindividual operating conditions can be considered as a priority. Whatthis means is that the operating conditions should have a greaterinfluence on the behavior of the entire system.

The advantage of the involute toothing typically used for gear pumps isthat the transmission ratio of the two rotational speeds remainsconstant during a rotation, which is a basic prerequisite for a constantvolume flow. In contrast, circular arc gear teeth have the disadvantagethat the transmission ratio of the rotational speed of the shafts variesperiodically and thus the flow of the conveyed medium pulsates. The useof the described invention with two controlled drive units makes itpossible for the first time to employ circular arc gear teeth without anunwanted pulsation of the current of the conveyed medium arising. Thus,with a sufficiently large backlash, the drive speeds of the shafts canbe corrected and compensated for with an opposing velocity profile, sothat circular arc gear teeth become possible with a constanttransmission ratio and hence constant volume flow.

As with the mentioned circular arc gear teeth other tooth forms are alsoconceivable. Merely the velocity profile needs to be adaptedaccordingly.

1. An arrangement including a gear pump, comprising a pump housing, twomeshing gear wheels contained in the pump housing and two shafts whichare operatively connected to the gear wheels and extend through the pumphousing, wherein the two shafts are each operatively connected torespective drive units, and wherein a coupling unit for compensation ofeccentricities between the drive unit and the respective shaft isarranged between each gear wheel and drive unit and wherein a rotaryencoder/sensor unit is arranged between the center of the gear wheel andthe center of the respective drive.
 2. Arrangement according to claim 1,wherein the rotary encoder/sensor unit is located in an axial regionwhich is defined by the center between the center of the gear wheel andthe center of the drive plus a deviation on both sides of at most 10% ofthe distance between the center of the gear wheel and the center of thedrive.
 3. Arrangement according to claim 2, wherein the rotaryspeed/sensor units are respectively arranged in the middle between therespective center of the gear wheel and the respective center of thedrive.
 4. Arrangement according to claim 1, wherein the rotaryencoders/sensor units feature a radial distance to a rotation axis ofthe respective shaft which is at least twice as large, as the outerradius of the gear wheels.
 5. Arrangement according to claim 1, whereinthe rotary encoders/sensor units are either optical or magnetic rotaryencoders/sensor units.
 6. Arrangement according to claim 1, wherein therotary encoders/sensor units are arranged such that a connecting linewhich runs through the corresponding rotary encoder/sensor unit andextends perpendicularly from the shaft encloses together with a planewhich centrally extends between the two rotation axes on a suction sidean angle in the range of 40° to 50°.
 7. Arrangement according to claim1, wherein each drive unit features a rotor and a stator, wherein therotor is axially moveable with respect to the stator.
 8. Arrangementaccording to claim 7, wherein the drive unit features on the far sidewith respect to the gear pump a differential bearing unit which radiallysupports the rotor of the drive unit.
 9. Arrangement according to claim8, wherein the rotor of the drive unit is connected to the respectiveshaft of the gear pump via the coupling unit, therewith radiallysupporting the rotor on the gear wheel side.
 10. Arrangement accordingto claim 1, wherein the coupling unit is a membrane coupling. 11.Arrangement according to claim 1, wherein a flange is arranged betweenthe pump housing and the stator the respective drive unit, wherein theflange features bores through which a cooling agent circulates foradjusting the temperature.
 12. Arrangement according to claim 1, whereinthe drive units are connectable to the respective shaft of the gear pumpfrom the far side with respect to the gear pump.
 13. Arrangementaccording to claim 12, wherein the connections between the drive unitsand the respective shafts of the gear pump are conical polygonconnections.
 14. Arrangement according to claim 1, wherein the driveunits are of the torque motor type.
 15. Arrangement according to claim1, wherein the one drive unit, the gear pump and the other drive unitare each contained in a temperature zone in which the temperatures areadjustable to specified values, and wherein isolation regions arepresent between neighboring temperature zones.
 16. Arrangement accordingto claim 1, wherein a current position of one gear wheel is determinablewith respect to the current position of the other gear wheel and whereinthe current position of the one gear wheel is continuously adjustablewith respect to the current position of the other gear wheel accordingto specified predefined operating conditions.